1. Field of the Invention
The present invention relates to a toroidal type continuously variable transmission which may be used as a transmission unit constituting a vehicular transmission or may be assembled as transmissions into various types of industrial machines.
2. Description of the Related Art
Study on the application of a toroidal type continuously variable transmission (as shown in FIGS. 1 and 2) into a vehicular transmission progresses. An example of the toroidal type continuously variable transmission is disclosed in Japanese Utility Model Unexamined Publication Sho. 62-71465.
In a conventional toroidal type continuously variable transmission shown in FIGS. 1 and 2, an input-side disk 2 is concentrically supported to an input shaft 1. An output shaft 3 is also disposed concentrically with an input shaft 1. An output-side disk 4 is fastened to the inner end of the output shaft 3. In the inside of a casing in which the toroidal type continuously variable transmission is stored, there are located a pair of trunnions 6, 6 at an intermediate position of the both disks 2, 4 along the axial direction thereof. The trunnions 6, 6 are swingable about their respective pivot shafts 5, 5 respectively disposed at position along an imaginary plane that is perpendicular to an imaginary line connecting the respective axes of the input and output shafts 1 and 3, and distanced from the intersection of the imaginary plane and imaginary line, as shown in FIG. 1. This physical relation is hereinafter referred to as "torsional relation".
Each of the trunnions 6, 6 located distant from the center axis of the input-side disk 2 and the output-side disk 4 is concentrically provided with each of the pivot shafts 5, 5 on the outer side surfaces of the two end portions thereof. The base end portions of displacement shafts 7, 7 are respectively supported in the central portions of the trunnions 6, 6 and if the trunnions 6, 6 are swung about the pivot shafts 5, 5 respectively, the inclination angles of the displacement shafts 7, 7 can be adjusted freely. On the peripheries of the two displacement shafts 7, 7 supported on the two trunnions 6, 6, there are rotatably supported a plurality of power rollers 8, 8 respectively. The power rollers 8, 8 are respectively interposed between the inner surfaces 2a and 4a, opposed to each other, of the input-side disk 2 and the output-side disk 4. The inner surfaces 2a and 4a are formed as concave surfaces which can be obtained by rotating an are having the pivot shaft 5 as a center thereof. And, the peripheral surfaces 8a, 8a of the power rollers 8, 8, which are formed as spherical-shaped convex surfaces are respectively in contact with the inner surfaces 2a and 4a.
Between the input shaft 1 and input-side disk 2, there is interposed a pressure device 9 of a loading cam type, while the input-side disk 2 is elastically pressed toward the output-side disk 4 by the pressure device 9. The pressure device 9 is composed of a cam plate 10 rotatable together with the input shaft 1, and a plurality of (for example, four pieces of) rollers 12, 12 which are respectively rollably held by a retainer 11.
On one side surface (in FIGS. 1 and 2, on the left side surface) of the cam plate 10, there is formed a drive-side cam face 13 being a curved surface which extends over the circumferential direction of the cam plate 10. And, on the outer surface (in FIGS. 1 and 2, on the right side surface) of the input-side disk 2, there is also formed a driven-side cam face 14 having a similar shape. The plurality of rollers 12, 12 are each rotatably supported about their respective shafts which extend in the radial direction with respect to the center of the input shaft 1.
The above-structured toroidal type continuously variable transmission operates in the following way. When the cam plate 10 is rotated with the rotation of the input shaft 1, the drive-side cam face 13 presses the plurality of rollers 12, 12 against the driven-side cam face 14 formed on the outer surface of the input-side disk 2. As a result of this, the input-side disk 2 is pressed against the plurality of power rollers 8, 8 and, at the same time the drive-side and driven-side cam faces 13 and 14 are pressed against the plurality of rollers 12, 12, so that the input-side disk 2 is rotated. The rotation of the input-side disk 2 is transmitted through the plurality of power rollers 8, 8 to the output-side disk 4, so that the output shaft 3 fastened to the output-side disk 4 is rotated.
Next, a description will be given of a case of changing of a rotational speed ratio (speed change ratio) of the input and output shafts 1 and 3. At first, when decelerating the rotational speed between the input shaft 1 and the output shaft 3, the trunnions 6, 6 are swung about the pivot shafts 5, 5 in a predetermined direction, respectively. Then, the displacement shafts 7, 7 are respectively inclined so that the peripheral surfaces 8a, 8a of the power rollers 8, 8, as shown in FIG. 1, can be respectively contacted with a near-center portion on the inner surface 2a of the input-side disk 2 and with a near-outer-periphery portion on the inner surface 4a of the output-side disk 4.
Also, on the other hand, when accelerating the rotational speed between the input and output shafts 1 and 3, the trunnions 6, 6 are respectively swung about the pivot shafts 5, 5 in the opposite direction to the predetermined direction. Then, the displacement shafts 7, 7 are respectively inclined so that the peripheral surfaces 8a, 8a of the power rollers 8, 8, as shown in FIG. 2, can be respectively contacted with a near-outer-periphery portion on the inner surface 2a of the input-side disk 2 and a near-center portion on the inner surface 4a of the output-side disk 4. When the inclination angles of the displacement shafts 7, 7 are set in the middle of the inclination angles shown in FIGS. 1 and 2, then there can be at obtained an intermediate transmission ratio between the input and output shafts 1 and 3.
A specific example of the toroidal type continuously variable transmission is shown in FIGS. 3 and 4. This transmission is disclosed in Japanese Utility Model Unexamined Publication No. Hei. 1-173552, recorded in a microfilm. As shown, an input-side disk 2 and an output-side disk 4 are rotatably supported around a cylindrical input shaft 15 with the aid of needle roller bearings 16, 16 inserted therebetween. A cam plate 10 is spline engaged with the outer peripheral surface of the end portion (in FIG. 3, the left end portion) of the input shaft 15 and is prevented, by a flange portion 17, from moving in a direction away from the input-side disk 2. Further the cam plate 10 and rollers 12, 12 constitute a pressure device 9 of a loading cam type. The pressure device 9, in accordance with the rotation of the input shaft 15, rotates the input-side disk 2 while it is pressing against the input-side disk 2 toward the output-side disk 4. An output gear 18 is coupled to the output-side disk 4 by means of keys 19, 19 so that the output-side disk 4 and the output gear 18 are synchronously rotated.
A pair of trunnions 6, 6, in particular, their respective two end portions thereof are supported on a pair of support plates 20, 20 in such a manner that they can be swung and can be displaced in the axial direction (in FIG. 3, in the front and back direction, or in FIG. 4, the horizontal directions) thereof. And, two displacement shafts 7, 7 are respectively supported in circular holes 21, 21 which are respectively formed in the middle portions of the pair of trunnions 6, 6. The two displacement shafts 7, 7 respectively include support shaft portions 22, 22 and pivot shaft portions 23, 23 which are extend in parallel to each other but are eccentric to each other. The support shaft portions 22, 22 are rotatably supported inside the circular holes 21, 21 through radial needle roller bearings 24, 24, respectively. Also, power rollers 8, 8 are rotatably supported in the peripheries of the pivotal support portions 23, 23 through another radial needle roller bearings 25, 25, respectively.
As shown in FIGS. 5 and 6 in detail, each of the radial needle roller bearings 25, 25 is constructed with a plurality of needle rollers 45, 45 and cage-like window type retainers 53 for holding rollably those needle rollers 45, 45. In this case, the outer circumferential surface of the pivot shaft portion 23 serves as a cylindrical inner raceway 54 of the radial needle roller bearing 25, and the inner circumferential surface of the power roller 8 serves as the outer raceway 55 of the radial needle roller bearing 25.
The pair of the displacement shafts 7, 7 are respectively disposed on 180 deg.-separated opposite sides with respect to the input shaft 15. Also, a direction, in which the pivot shaft portions 23, 23 of the displacement shafts 7, 7 are eccentric to the support shaft portions 22, 22, is set as the same direction with respect to the rotation direction of the input- and output-side disks 2 and 4. Also, the eccentric direction is set almost at right angles to the direction in which the input shaft 15 is disposed. Therefore, the power rollers 8, 8 are supported in such a manner that they can be somewhat displaced in the disposing direction of the input shaft 15. As a result, even when, due to accumulation of the dimensional tolerance of the components parts, the input- and output-side disks 2 and 4 are displaced from the trunnions 6, 6 in the axial direction of the input shaft 15 (in FIG. 3, the horizontal direction, or in FIG. 4, front-back direction) to some degree, adequate contact of the inner surface 2a and the inner surface 4a of the disks 2 and 4 with the peripheral surfaces 8a of the power rollers 8 is secured. Further, when the component parts are deformed by large loads imparted thereto in a transmission state of the rotational force, and as a result of the deformation, even if the power rollers 8, 8 are likely to displace in the axial direction of the input shaft 15, this displacement of the power rollers 8, 8 may be absorbed without applying excessive force to the component parts.
Also, between the outer surfaces of the power rollers 8, 8 and the inner surfaces of the middle portions of the trunnions 6, 6, there are interposed thrust ball bearings 26, 26 and thrust needle roller bearings 27 are disposed in this order from the outer surfaces of the power rollers 8. The thrust ball bearing 26, 26 are respectively used to allow the power rollers 8, 8 to rotate while supporting the load applied to the power rollers 8, 8 in the thrust direction. The thrust ball bearings 26, 26 are respectively composed of a plurality of balls 56, 56, annular-shaped retainers 57, 57 for rollably holding the balls 56, 56 therein, and annular-shaped outer races 28, 28. The inner raceways of the thrust ball bearings 26, 26 are respectively formed on the outer surfaces of the power rollers 8, 8, whereas the outer raceways thereof are respectively formed on the inner surfaces of the outer races 28, 28.
Each of the thrust needle roller bearings 27, 27 is composed of a race 58, a retainer 59 and needle rollers 60, 60. The race 58 and retainer 59 are combined together in such a manner that they can be somewhat displaced in the rotation direction. The thrust needle roller bearings 27, 27 interpose the races 58, 58 between the inner surfaces of the trunnions 6, 6 and the outer surfaces of the outer races 28, 28 in a state that the races 58, 58 are contacted with the inner surfaces of the trunnions 6, 6. The thrust needle roller bearings 27, 27 allow the pivot shaft portions 23, 23 and the races 28, 28 to rotate about the support shaft portions 22, 22 while receiving a thrust load applied to the outer races 28, 28.
Drive rods 29, 29 are respectively coupled to one end portions (left end in FIG. 4) of the trunnions 6, 6. And, drive pistons 30, 30 are respectively firmly coupled to the outer surface of the middle position of the drive rods 29, 29. The drive pistons 30, 30 are oil-tightly disposed within drive cylinders 31, 31. An amount of displacement of each of the trunnions 6, 6, which is caused by supplying oil into and discharging it from each of the drive cylinders 31, 31, is detected by a precess cam (not shown) fixed to the other end portions of the trunnions 6, 6.
A lubricating-oil supplying device as shown in FIG. 7 is provided in the insides of the drive rod 29, the trunnion 6 and the displacement shaft 7. The lubricating-oil supplying device feeds a sufficient amount of lubricating oil into the bearings 25 and 26 in order to secure the durability of the radial needle roller bearing 25 and the thrust ball bearing 26. The lubricating-oil supplying device is composed of a feeding-side oil-supply passage 42 provided in the insides of the drive rod 29 and the trunnion 6, oil-feedholes 43, 43 formed in the outer race 28 of the thrust ball bearing 26, and a receiving-side oil-supply passage 44 provided in the inside of the pivot shaft portion 23, which constitutes the first half of the displacement shaft 7. When the toroidal type continuously variable transmission is in operation, the lubricating-oil supplying device feeds lubricating oil into the feeding-side oil-supply passage 42 with the aid of a pump (not shown) assembled into the transmission, to thereby lubricate the bearings 25 and 26.
In the thus constructed toroidal type continuously variable transmission, a rotation of the input shaft 15 is transmitted to the input-side disk 2 through the pressure device 9. A rotation of the input-side disk 2 is transmitted through the pair of power rollers 8, 8 to the output-side disk 4, and a rotation of the output-side disk 4 is output from the output gear 18. To change the rotational speed change ratio between the input shaft 15 and the output gear 18, the pair of drive pistons 30, 30 are displaced in the opposite directions to each other. In accordance with the displacement of the drive pistons 30, 30, the pair of trunnions 6, 6 displace in the opposite directions, so that the lower power roller 8 disposed in the downside of FIG. 4 displaces to the right, while at the same time the upper power roller 8 disposed in the upside of FIG. 4 displaces to the left. Accordingly, the direction of forces in the tangential direction which act on contact positions where the peripheral surfaces 8a, 8a of the power rollers 8, 8 are in contact with the inner surface 2a of the input-side disk 2 and the inner surface 4a of the output-side disk 4, is changed. In accordance with the changing of the direction of the forces, the trunnions 6, 6 are swung about the pivot shafts 5, 5 which are supported by the support plates 20, 20 in the opposite directions to each other. As a result, as shown in FIGS. 1 and 2, the contact positions where the peripheral surfaces 8a, 8a of the power rollers 8, 8 are in contact with the inner surface 2a and the inner surface 4a of the input-and output-side disks 2 and 4 are shifted, whereby the rotational speed change ratio between the input shaft 15 and the output gear 18 is changed. The control of the rotational speed change ratio to a desired value is conducted in a manner that the amounts of the displacements of the trunnions 6, 6 in the axial directions of the pivot shafts 5, 5, which are detected by the precess cam, is adjusted by adjusting the amounts of the pressurized oil charged to and discharged from the drive cylinders 31, 31.
When the rotational force is transmitted between the input shaft 15 and the output gear 18, based on the elastic deformation of the component parts, the power rollers 8, 8 are displaced in the axial direction of the input shaft 15. As a result, the displacement shafts 7, 7 which pivotally support the power rollers 8 are slightly turned about the support shaft portions 22, respectively. Due to the turning of the displacement shafts 7, 7, the outer surfaces of the outer races 28, 28 of the thrust ball bearings 26, 26 are displaced relative to the inner surfaces of the trunnions 6, 6. A force required 5 for the relative displacement is small because the thrust needle roller bearings 27 are present between the outer surfaces of the races 28, 28 and the inner surfaces of the trunnions 6, 6. This fact implies that a force to change an inclination angle of each of the displacement shafts 7, 7 is small.
Turning now to FIGS. 8 and 9, there are shown toroidal type continuously variable transmissions increased in their transmissible torque. As shown, a couple of input disks 2A and 2B and a couple of output disks 4, 4 are arranged side by side around an input shaft 15a in the power transmission direction. In either structure (FIGS. 8 and 9), an output gear 18a is disposed in a middle portion of the input shaft 15a to be rotatably supported around the input shaft 15a. The output disks 4,4 are spline-engaged to both ends of a cylindrical 20 sleeve 32 provided in the central portion of the output gear 18a. Needle roller bearings 16, 16 are respectively provided between the inner circumferential surfaces of the output disks 4, 4 and the outer circumferential surface of the input shaft 15a. With provision of the needle roller bearings 16, the 25 output disks 4, 4 are supported around the input shaft 15a so as to be rotatable about the input shaft 15a and movable in the axial direction of the input shaft 15a. The input disks 2A and 2B are supported at both ends of the input shaft 15a while being rotatable together with the input shaft 15a. The input shaft 15a is rotatable driven by a drive shaft 33 through the pressure device 9 of the loading cam type. There is provided a radial bearing 34, such as a sliding bearing or a needle roller bearing, is disposed between the outer circumferential surface of the tip end (right end of in FIGS. 8 and 9) of the drive shaft 33 and the inner circumferential surface of the base end (left end in FIGS. 8 and 9) of the input shaft 15a. Therefore, the drive shaft 33 and the input shaft 15a are concentrically combined with each other such that those shafts are slightly movable in the rotational direction.
The rear surface of input-side disk 2A (located on the right side in FIGS. 8 and 9) is thrust against a loading nut 35 directly (in the structure shown in FIG. 9) or with a coned disk spring 36 having large resilience being interposed therebetween (in the structure shown in FIG. 8), to thereby substantially prevent the displacement of the input-side disk 2A in the axial directions (horizontal directions in FIGS. 8 and 9) of the input shaft 15a. On the other hand, the input-side disk 2B facing the cam plate 10 is supported to be movable in the axial direction of the input shaft 15a with the aid of a ball spline 37. A coned disk spring 38 and a thrust needle roller bearing 39 are serially disposed between the rear surface (right-side surface in FIGS. 8 and 9) of the input-side disk 2B and the front surface (right-side surface in FIGS. 8 and 9) of the cam plate 10. The coned disk spring 38 functions so as to impart pre-load to contact portions where the inner surfaces 2a of the input-side disks 2A and 2B and the inner surface 4a of the output-side disk 4 are in contact with the peripheral surfaces 8a, 8a of the power rollers 8, 8. The thrust needle roller bearing 39 allows the input-side disk 2B to rotate relative to the cam plate 10 when the pressure device 9 operates.
In the structure of FIG. 8, the output gear 18a is rotatably supported while the axial displacement thereof being prevented, on a partitioning wall 40 provided inside of the housing, by a pair of ball bearings 41, 41 of the angular type. In the structure of FIG. 9, the output gear 18a is axially displaceable. In the toroidal type continuously variable transmission of the double cavity type in which the couple of input-side disks 2A and 2B and the couple of output-side disks 4,4 are arranged side by side in the power transmission direction, as shown in FIGS. 8 and 9, one of the input-side disks 2A and 2B, which faces the cam plate 10 or both of them is or are axially movable with respect to the input shaft 15a by means of the ball spline 37, 37a. The reason for this is that the transmission structure is designed so as to allow the input-side disks 2A and 2B to displace in the axial directions of the input shaft 15a, while securing the synchronous rotations of the input-side disks 2A and 2B, based on the elastic deformation of the related component parts due to operations of the pressure device 9.
The ball spline 37 and ball spline 37a include inner-diameter ball-spline grooves 62 formed in the inner circumferential surfaces of the input-side disks 2A and 2B, outer-diameter ball-spline grooves 63 formed in the outer circumferential surfaces of the intermediate portion of the input shaft 15a, and a plurality of balls 64, 64 rollably provided between the inner-diameter ball-spline grooves 62 and the outer-diameter ball-spline grooves 63. As for the ball spline 37 for supporting the input-side disk 2B located closer to the pressure device 9, a stopper ring 66 is retained in a stopper groove 65 formed in a portion of the inner circumferential surface of the input-side disk 2B, which is closer to the inner surface 2a thereof, to thereby limit the balls 64, 64 in displacing toward the inner surface 2a of the input-side disk 2B. Further, it prevents the balls 64, 64 from slipping off from between the inner-diameter ball-spline grooves 62 and the outer-diameter ball-spline grooves 63. As for the ball spline 37a for supporting the input-side disk 2A located apart from the pressure device 9 in the transmission structure of FIG. 8, a stopper ring 66a is retained in a stopper groove 65a formed in the outer circumferential surface (a portion thereof closer to the left end in FIG. 8) of the input shaft 15a, to thereby limiting the balls 64, 64 in displacing toward the inner surface 2a of the input-side disk 2A.
In the known or proposed toroidal type continuously variable transmission, less consideration is given to the eccentric quantities of the displacement shafts 7, 7 for supporting respectively the power rollers 8, 8 on the inner surfaces of the intermediate portions of the trunnions 6, 6. The support shaft portion 22, 22 and the pivot shaft portion 23, 23 are parallel to each other, but the former is eccentric from the latter, viz., their centers are not coincident with each other (FIGS. 13, 24 and 25). Little qualitative consideration has been made on an eccentric quantity L.sub.7 present between the support shaft portion and the pivot shaft portion 23, 23. The study by inventor(s) on the toroidal type continuously variable transmission showed the following fact: To extract desired performances of the toroidal type continuously variable transmission, it is essential to place the eccentric quantity L.sub.7 within a proper range of eccentric quantity values. This fact will be described by use a case where the toroidal type continuously variable transmission of the double cavity type as shown in FIG. 10 is in a maximum deceleration state where trouble occurrence is most frequent.
When the eccentric quantity L.sub.7 is excessively small, the speed change ratio of the toroidal type continuously variable transmission shifts from a desired speed change ratio for the following reason. To absorb the dimensional tolerance of the component parts and the elastic deformations of those parts during the power transmission, the pivot shaft portion 23 constituting each displacement shafts 7 revolves around the support shaft portion 22. For example, at the time of the transmission of power, a thrust load that is generated by the pressure device 9 thrusts the output-side disk 4. The output-side disk 4 is elastically displaced from a position (dot chain line in FIG. 11) to another position (solid line in FIG. 11), and the input-side disk 2B is displaced toward the output-side disk 4 (right side in FIG. 11). In accordance with the displacement, the power roller 8 held between the inner surface 2a of the input-side disk 2B and the inner surface 4a of the output-side disk 4 moves in the axial direction (referred to as an x-direction, for ease of explanation) of the input shaft 15a. With the movement, the trunnion 6, the displacement shaft 7 and the power roller 8 changes from their disposition of FIG. 12A to another disposition of FIG. 12B. The change of the disposition of those components results from the revolution of the pivot shaft portion 23 with respect to the support shaft portion 22. Therefore, the pivot shaft portion 23 and the power roller 8 move also in the axial direction (referred to as a y-direction, for ease of explanation) of the pivot shafts 5, 5 which pivotally supports the trunnion 6 as well as in the x-direction, as shown in FIGS. 13A and 13B.
The movement of the pivot shaft portion 23 and the power roller 8 in the y-direction, as seen from the above description, is the same as the operation of them in a case where the trunnions 6 are displaced in the axial direction of the pivot shafts 5, 5 by moving forward and backward the drive rods 29 (see FIG. 4) to change an inclination angle of the power roller 8 for the purpose of changing the rotational speed change ratio of the input-side disk 2B and the output-side disk 4. Accordingly, when the power roller 8 displaces in the x-direction, on the basis of the displacement in the y-direction which is simultaneously applied, the power roller 8 is displaced by a distance corresponding to the displacement in the y-direction caused by the revolution, although the trunnion 6 per se does not displace in the y-direction. When a degree of speed change (speed change quantity), which is caused by such a displacement of the power roller is small, no problem arises. When it is too much large, the speed change ratio cannot be controlled as desired.
To control the speed change ratio of the toroidal type continuously variable transmission, a controller decides a target speed change ratio based on a signal representative of throttle-valve position, engine speed, or running speed; an instruction signal indicative of the target speed change ratio is applied to a related electric motor; and controls the switching of a hydraulic-pressure control valve, and thus operates the drive pistons 30 (FIG. 4). And, the contact positions where the peripheral surfaces 8a of the power rollers 8 are in contact with the inner surface 2a of the input-side disk 2 (2A, 2B) and the inner surface 4a of the output-side disk 4 are shifted to other positions, so as to change the inclination angles of the power rollers 8. However, where a quantity y8 of a displacement of the power roller 8 in the y-direction, caused by the revolution motion, is increased, another action not caused by the signals stated above exists in addition to the action for the changing of the speed change ratio, which is caused by the drive pistons 30, 30. Therefore, the toroidal type continuously variable transmission changes its speed change ratio. Further, an actual speed change ratio is greatly deviated from the target one, and the toroidal type continuously variable transmission operates in a region out of an optimum region of its characteristic where the fuel consumption by the engine is efficient and the output power of the engine is high. This situation should be avoided.
In the conventional technique, it is considered that the preferable way to suppress the y-directional movement of the power roller 8, which is produced when the power roller 8 is moved in the x-direction is to secure the eccentric quantity L.sub.7 of the support shaft portions 22, 22 from the pivot shaft portions 23, 23 as large as possible. Further, it is recognized that where the eccentric quantity L.sub.7 is excessively large, a cross sectional area of the joint portion where the support shaft portions 22, 22 and the pivot shafts portions 23, 23 are jointed together is small, and as a result, a stress generated in the joint portion is great and in this condition it is very difficult to secure a satisfactory durability of the displacement shafts 7, 7. Therefore, the designer considers that the eccentric quantity L.sub.7 has certain values of the upper limit, and they determine the eccentric quantity L.sub.7 on the basis of the best balance between the securing of the durability of the displacement shaft and the suppressing of the y-directional component.
As described above, the conventional design of the eccentric quantity L.sub.7 between the support shaft portions 22, 22 and the pivot shaft portions 23, 23 constituting the displacement shafts 7, 7 is not based on definite rules constructed in consideration with the performance on the speed-ratio change of the toroidal type continuously variable transmission. The inventor (s) discovered that there is a specific correlation between the eccentric quantity L.sub.7 and the speed-ration change performance of the toroidal type continuously variable transmission, and that the eccentric quantity L.sub.7 with a specific range, provides a satisfactory speed-ratio change performance.
Further, in designing the conventional toroidal type continuously variable transmission, any special consideration has been given to the surface natures of the displacement shafts 7 which are used for supporting the power rollers 8, 8 on the trunnions 6, 6 in rotatable and displaceable fashion. Therefore, a satisfactory durability of the transmission is not always guaranteed where the transmission is used under hard conditions. The reason for this will be described with reference to FIGS. 14 through 17. When the toroidal type continuously variable transmission is in operation, the power roller 8 is strongly compressed between the input-side disk 2 and the output-side disk 4 as shown in FIG. 14. Accordingly, the center hole of the power roller 8 is deformed to be elliptical as exaggeratedly illustrated in FIG. 15. In this state, the pivot shaft portion 23 of the displacement shaft 7 is strongly thrust in the directions in which the input-side disk 2 and the output-side disk 4 are arranged.
When the power roller 8 is strongly compressed between the input-side disk 2 and the output-side disk 4, a large force thrusts the power roller 8 outwardly in the radial directions of the input-side disk 2 and the output-side disk 4 when viewed in cross section, since the peripheral surfaces 8a of the power roller 8 is engaged with the inner surface 2a of the input-side disk 2 and the inner surface 4a of the output-side disk 4. Due to the thrust forces, the trunnion 6 supporting the power roller 8 on its inner surface is elastically deformed from the configuration shown in FIG. 16A to the configuration shown in FIG. 16B. Since the support shaft portion 22 of the displacement shaft 7 is somewhat offset from the center of the trunnion 6, the displacement shaft 7 is inclined by the elastic deformation of the trunnion 6. The inclination of the displacement shaft 7 leads to partial contact of the outer circumferential surface of the pivot shaft portion 23 of the displacement shaft 7 with the needle rollers 45, 45 constituting the radial needle roller bearing 25. More particularly, as shown by oblique lattices in FIG. 17, rolling surfaces of the needle rollers 45, 45 are strongly pressed against the outer circumferential surface of the pivot shaft portion 23.
The partial contact by the elastic deformation of the power roller 8 and the partial contact by the inclination of the displacement shaft 7 are summed, so that load regions as indicated by oblique lattices in FIG. 18 appear in the pivot shaft portions 23. In those load regions, large area pressure is applied from the rolling surfaces of the needle rollers 45, 45 to the outer circumferential surfaces of the pivot shaft portions 23. The surface roughness of the rolling surface (the inner and outer raceway portions being in contact with the rolling surfaces of the needle rollers 45, 45) of a general radial needle roller bearing, used in a high speed region of 10,000 rpm or higher, is about 0.4 .mu.mRa. However, since the rolling surfaces of the needle rollers 45, 45 are strongly contacted with the outer circumference surface of the pivot shaft portion 23 in the above load regions, an oil film is hard to be formed on the contact portions when the surface roughness of the outer circumference surface is about 0.4 .mu.mRa.
In the portions on which large area pressure exerts, a large amount heat is generated according to the operation of the toroidal type continuously variable transmission. Those portions are also located close to traction portions where the peripheral surfaces 8a of the power roller 8 are in contact with the inner surface 2a of the input-side disk 2 and the inner surface 4a of the output-side disk 4. Elevation of temperature caused by the heat generated in the traction portions is great. Accordingly, the heat-resistance of those portions receiving the large area pressure needs to be secured for securing a satisfactory durability of the displacement shaft 7.
In addition, in the conventional toroidal type continuously variable transmission, the radial needle roller bearings 25 which rotatably support the power rollers 8 around the pivot shaft portions 23 of the displacement shafts 7, respectively, are not always satisfactory in their durability. The reason for this will be described hereunder.
Where the toroidal type continuously variable transmission is used for a transmission unit of a motor vehicle, an automotive power that is output from the engine to the input shafts 15, 15a is transmitted to the output-side disk 4, through the input-side disk 2, 2A, 2B and the power rollers 8, 8. The toroidal type continuously variable transmission may be considered in the form of the radial needle roller bearings 25, which support the power rollers 8, 8 around the pivot shaft portions 23, respectively. In this case, it is operated in an outer race rotating mode in which the power roller 8 having the outer raceway 55 revolves. A load applied to the thus radial needle roller bearing 25 is a radial component of a force, that is, a traction force, applied to the traction portions of the power roller 8 supported by the radial needle roller bearing 25, viz., the contact portions where the inner surfaces 2a of the input disks 2A and 2B and the inner surface 4a of the output-side disk 4 are in contact with the peripheral surfaces 8a of the power rollers 8.
The radial load applied to the radial needle roller bearing 25 varies depending on the output power (in particular torque) of the engine and a changing state of the speed change ratio of the toroidal type continuously variable transmission. In the case of a normal aspiration engine of the displacement volume of 2,000 to 3,000 cc, the radial load is approximately 500 to 700 kgf (5000 to 700N) under the condition that the toroidal type continuously variable transmission is in a maximum deceleration state and a maximum torque input state. In the case of the natural aspiration engine of 800 cc to 1500 cc in displacement volume, it is approximately 200 to 400 kgf (2000 to 4000 N) under the same condition as above.
The radial needle roller bearing 25 is capable of sufficiently enduring such a radial load if it is under a general load loading condition. However, the power roller 8, which functions as the outer race of the radial needle roller bearing 25, is repeatedly elastically deformed due to loads from the inner surface 2a of the input-side disk 2, 2A, 2B and the inner surface 4a of the output-side disk 4. Therefore, an excessive area pressure acts on a part of the rolling contact surface, and the durability of the power roller 8 is possibly lost. This will be described with reference to FIGS. 19 to 22.
When the toroidal type continuously variable transmission is in operation, loads indicated by an arrow a in FIGS. 19 to 20 are imparted to two opposed positions on each of the power rollers 8, 8 from the inner surface 2a of the input-side disk 2, 2A, 2B and the inner surface 4a of the output-side disk 4. As seen from FIGS. 19 to 20, those loads are directed toward the positions on the power rollers 8, 8 closer to the trunnions 6, 6. When the loads directed to the arrow .alpha. are increased in value, the inside diameters of the power rollers 8, 8 are elastically deformed as exaggeratedly shown in FIG. 21, the outer raceway 55 is deformed to be elliptical in cross section as exaggeratedly illustrated in FIG. 22. In this case, the amount of deformation of the outer raceway 55 is not caused in the axial direction of the radial needle roller bearing 25 and increases in quantity toward the trunnions 6, 6 with respect to the radial direction thereof. At a specific portion in the circumferential direction of the outer raceway 55, the elastic deformation inwardly in the radial direction thereof is conducted two times during one turn of each power roller 8.
As the result of the elastic deformation of the outer raceway 55, the distance between the inner raceway 54 and the outer raceway 55 of the radial needle roller bearing 25 becomes narrower at two opposite positions in the radial direction where it faces the inner surface 2a of the input-side disk 2, 2A, 2B and the inner surface 4a of the output-side disk 4, and it is close to the trunnion 6. At those positions, the needle rollers 45, 45 of the radial needle roller bearing 25 are forcibly compressed between the inner raceway 54 and the outer raceway 55. As a result, an excessive area pressure, which is due to an edge load, is applied to parts of the inner raceway 54 and the outer raceway 55, which face the ends of the needle rollers 45, 45 (when axially viewed). The excessive area pressure causes early flaking-off on those portions.
When the portions are damaged by such pressure-flaking, sound and vibration generated at the radial needle roller bearing 25 become large. As a result, sounds and vibrations generated by not only the toroidal type continuously variable transmission having the radial needle roller bearings assembled thereinto but also the transmission unit having the toroidal type continuously variable transmission, are increased. This adversely affects the drive feeling of the vehicle having the transmission unit. Further, when flakes separated from the traces enter into the traction portion transmitting the automotive power, the area pressure excessively increases thereat. This possibly causes the damages such as the flaking in the early stage in the inner surface 2a of the input-side disk 2, 2A, 2B and the inner surface 4a of the output-side disk 4, and the peripheral surfaces 8a, 8a of the power rollers 8, 8, which form the traction portion. Moreover, the strainer and the filters may be clogged with the flakes thus caused. This results in reduction of the discharge amount of the pump for supplying the lubricating oil, poor lubricating, and reduction of lifetime of other parts.